Our invention relates generally to parallel shaft transmissions for automotive vehicles, particularly multiple ratio transmissions having a mainshaft adapted to be driven by the vehicle engine crankshaft axis and a countershaft arranged in parallel disposition with respect to the mainshaft.
A typical transmission of this kind may include multiple gears of different pitch diameters journalled on the common axis of a torque input shaft and an output shaft. The gears engage companion gears mounted on the axis of a countershaft. Such a typical arrangement is shown in prior art reference U.S. Pat. No. 3,618,416, where torque delivery gears are journalled on a mainshaft and are arranged in constant mesh with countershaft gear elements. Synchronizer clutch assemblies are provided in the design of the '416 patent for selectively connecting each of the gears to the mainshaft to establish selectively the various drive ratios. The gears continuously rotate with respect to the mainshaft axis, even though only one of them is conditioned at any given time by its respective synchronizer clutch for torque delivery.
Another example of a parallel-shaft, constant-mesh, shiftable transmission with synchronizers mounted on the axis of the mainshaft is shown in prior art U.S. Pat. No. 3,994,182. This patent is distinguishable from the transmission of the '416 patent primarily by the construction of the countershaft gearing. In the case of the '416 patent, the countershaft includes independent cluster gears that rotate with and are connected to the countershaft itself; whereas the design of the '182 patent includes integral cluster gearing which is journalled on a fixed or "dead" mounting shaft. The mounting shaft, since it does not rotate, does not create inertia torque that is added to the inertia torque of the rotating countershaft gearing.
Synchronizer clutch assemblies for a double-shaft, constant-mesh, shiftable transmission may also be mounted on the axis of the countershaft rather than the common axis of the input and output shafts. An example of this construction is shown in prior art U.S. Pat. No. 4,757,726 where a synchronizer for the over-drive ratio is mounted on the countershaft and the synchronizers for the direct-drive ratio and three of the other forward drive ratios are mounted on the axis of the mainshaft.
Arrangements of this kind exhibit a tendency for the gears that are not functioning at any given instant in the torque flow path to produce an undesirable gear rattle. Gear rattle is induced by transient engine accelerations that are cyclic in nature. All driven rotary components of the transmission and the driveline components in the torque flow path downstream of the engine are affected by these angular accelerations. The extent to which they are effected depends upon the individual inertias of these components.
The magnitude of these acceleration effects depends also partly on the transmission gear ratio and the location of the individual components in the transmission or elsewhere in the driveline.
Gear rattle is discernible by the impacting of the gear teeth at each gear mesh in those torque flow paths that are inactive when the transmission is operated in any given ratio or when the transmission is operated in neutral.
Gear rattle is induced when the inertia torque of a component exceeds the drag torque acting on that component. If at any given gear mesh the inertia torques are sufficient to overcome the effect of frictional drag, including the viscous shear drag of the lubricating oil, a gear rattle is likely. Inertia torque is the product of angular acceleration and the inertia of the driven component.
During a ratio change in a synchronized transmission, a synchronizer sleeve in the synchronizer assembly exerts a force on a synchronizer blocker ring as chamfers on the blocker ring teeth engage chamfers on the internal teeth of the sleeve. The blocker ring in turn exerts a torque on the gear involved in the ratio change, causing the gear to accelerate until it approaches the angular velocity of the shaft on which it is mounted. The synchronizer blocker ring inhibits completion of the shifting of the sleeve until synchronization is achieved.
The gears that are in mesh with the gear being synchronized also must be accelerated or decelerated, as well as the other rotating parts of the transmission that rotate with the gear being synchronized. A reduced inertia for this rotating mass results in a reduced synchronizing force for effecting a ratio change and a shorter synchronizing time. This improves the shiftability.
We are aware of various prior attempts to overcome the gear rattle problem. Examples of prior art designs include transmissions that have brake members to provide a frictional drag on the gears. Thus, when the gears are inactive, a gear rattle will not occur as long as the frictional drag is above a threshold inertia torque at the gear mesh. This is an undesirable solution to the gear rattle problem because it contributes to the mechanical inefficiency in the transmission. It also has an undesirable effect on shiftability. A greater shift effort is required in the gear shift linkage if a parasitic drag is induced by the anti-rattle elements. This parasitic drag, when added to the effective inertia of the clutch and the related rotating components of the transmission, must be overcome by the operator as a shifting force is applied through the shift linkage to the synchronizer clutch assemblies.
We are aware also of other prior art attempts to eliminate the gear rattle problem by introducing a so-called scissor gear tooth arrangement in which one element of the meshing gearset is divided into two parts that are spring loaded torsionally, thereby eliminating a backlash condition at the conjugate meshing of the gear teeth. Still another example of an attempt to eliminate gear rattle is shown in prior art U.S. Pat. No. 4,811,615 where a torsional load is applied to adjacent elements of a countershaft gear so that the spring torque applied is of greater magnitude than the inertia torque at the meshing gear teeth. In this instance, the torsional spring load is capable of overcoming the torsional acceleration of the individual gears, provided the inertia torque does not exceed the spring torque. Aside from the functional limitations of these design arrangements, they add to the design complexity and cost and increase the space requirements for the gearing.